1. Field of the Invention
This invention relates generally to the field of internal combustion engines and more particularly to a multicylinder internal combustion engine having a wobbler connected to the pistons and positionable on a dual angled crank for variation of the stroke while maintaining constant compression ratio.
2. Description of the Related Art
The internal combustion engine in its two current forms (spark ignited gasoline and compression ignition diesel) is poorly suited to efficiently power personal transportation. Automobiles and customer's expectations have evolved such that the engine's maximum power is far beyond what is normally used and the engine operates at 5-25% load most of its life.
The diesel engine is more efficient at part load than the gasoline engine; however the 2000+ bar fuel injection system, variable geometry turbocharger(s) system, and emissions after-treatment have made the modern passenger vehicle diesel engine quite expensive. High levels of EGR to reduce NOX, SCR catalysts, and diesel particulate filters all negatively impact the engine efficiency. In addition, the price of diesel fuel relative to gasoline often negates any cost savings from the inherent higher efficiency of the engine for the end user.
The gasoline engine is still relatively inexpensive, and due to the 3 way catalyst, it is a very clean engine and full load efficiency can reach 35%. Unfortunately, the part load performance is extremely poor, with thermal efficiency dipping into single digits during much of its operation and seldom reaching 20% or higher. An idling gasoline engine with the vehicle running its air conditioning, power steering, and entertainment system has significant energy losses from throttling, engine friction accessory parasitic losses and cooling losses and uses fuel at a very high rate.
FIG. 1 shows the basic Pressure/Volume (PV) diagram for a gasoline engine. Power is controlled by throttling the air system. The engine runs at a nearly constant air/fuel ratio, therefore the engine's air system must be restricted to reduce the fuel flow and thus the power. At low load operation, the intake manifold of the engine is reduced to a very low pressure, while the exhaust manifold remains above atmospheric (plus muffler, catalyst, and pipe restrictions). This results in a severe pumping loss. The difference in pressure from the engine's intake to exhaust due to throttling is a direct loss in Brake Mean Effective Pressure (BMEP) and is commonly referred to as “negative pumping loop work” as shown in FIG. 2.
Perhaps an even larger impact than the direct pumping work, the engine system's Compression Ratio (CR) and Expansion Ratio (ER) are reduced. Thermodynamic cycles (such as the Otto cycle) are often used to characterize various engine types, however these cycles are different than the actual engine mechanical cycle that is seen on a PV diagram. The fact that a PV diagram of the actual engine cycle looks similar to the thermodynamic cycle adds to the confusion. In the thermodynamic cycle, the compression ratio and expansion ratio are characterized on the X-axis, i.e. Volume Ratio. This does not account for the induction and exhaust processes. If it did, then those processes would have to happen instantaneously at top dead center and bottom dead center with no pressure drop and the connecting reservoir would have to be at the same pressure and temperature, i.e. no throttling on the inlet or back pressure on the exhaust.
The theoretical volume ratio of an engine is not indicative of real efficiency when intake valves and exhaust valves open at various times during the compression and expansion processes, and an engine can throttled down to 0.1 bar in the intake manifold while the exhaust manifold is greater than 1 bar. The limitations of using thermodynamic cycles to draw conclusions about real engines are particularly severe when considering part load operation of a throttled gasoline engine.
In the real engine, power is produced by the expansion of the hot gas; the compression of the cold air is a necessary parasitic loss. These powers, one positive and one negative, are a function of the compression ratio and expansion ratio as calculated on the Y-axis (Pressure Ratio).
In a closed thermodynamic cycle, there is a fixed relationship between P and V, i.e. PV=mRT. In an open cycle engine with valves, this relationship does not exist, therefore the thermodynamic cycles are not really simplified representations of real operating cycles.
The efficiency of the simplified thermodynamic cycle that represents this engine, the Otto cycle, is a function of compression ratioη=1−1/CR(γ−1) where                η engine efficiency        CR compression ratio        γ ratio of specific heats of air        
Looking at a highly throttled operating point as shown in FIG. 2, the volume compression in the cylinder is the same ratio, but since the intake manifold pressure is very low, the absolute pressure of the gas that is compressed is much lower. The power of the engine is obtained by the expansion stroke. If the outlet pressure of the expansion stroke remains fixed at slightly above atmospheric, then the lower pressure from compression will result in a lower expansion ratio. This reduces the power and efficiency of the engine. As an example, if a 10/1 compression ratio engine is throttled from 1 Bar to 0.5 Bar in the intake manifold, the true CR to put into the efficiency calculation is a CR of 5/1. While there may be some academic debate that the compression ratio is determined by the displacement of the engine and the minimum combustion volume rather than the pressures at the beginning and end of compression, there is no debate that the expansion ratio of the engine will be less when it is throttled, which is clearly shown in FIG. 2. Therefore, the engine power is reduced not just from the reduction of air mass flow (and thus fuel flow), but from having a lower expansion ratio. The reduction of mass flow through the engine reduces the power—the reduction of the expansion ratio lowers the power and the thermal efficiency.
The engine friction comes from a great number of components; rings, pistons, rod bearings, crankshaft bearings, and the valve train (cam bearings, tappets, valves, and gears or chain). The oil pump parasitic loss can be book kept here as well. As known in the art, the rings and pistons account for about ½ of the friction loss as shown in FIG. 3 (TMechE Review on Tribology. “Tribology—Motoring into the 21st Century” by Chris Taylor, 2003). The friction of the engine is mainly a function of the engine speed and not of load. The Shinn-Flynn friction model for IC engines shows this as does FIG. 4 (Internal Combustion Engine Handbook, Edited by Basshuysen & Shafer, 2004 SAE International):FMEP=a+(b*Pcyl)+(c*v2)                where,        FMEP Friction Mean Effective Pressure        a constant part of friction (0.3-0.5 bar)        b coefficient for contribution due to cylinder maximum pressure (0.004-0.006)        Pcyl peak cylinder pressure (bar)        c coefficient for contribution due to piston (0.0006-0.0012)        v mean piston speed (m/sec).        
Thus an engine at 2000 rpm accelerating at full load has essentially the same friction loss as an engine at 2000 rpm at 10% load (cruising at moderate speed). If the output of the engine is 40 kW and the friction loss is 2 kW, then it's a 5% effect. If the output of the engine is at 4 kW, and the friction loss is still 2 kW, then it's a 50% effect. This fact shows why cylinder deactivation schemes bring only marginal benefit. The pistons, rings, crank, and valve train are still contributing very nearly the same parasitic loss as they would be if all the cylinders were activated.
It is therefore desirable to provide a gasoline internal combustion engine which reduces losses from throttling and engine friction.